1. Field of the Invention
The present invention relates generally to a flow control valve adapted for being incorporated in a variable displacement refrigerant compressor. More particularly, the present invention relates to a variable displacement refrigerant compressor accommodating therein a flow control valve which permits the compressor to be incorporated in a refrigerating system for a vehicle climate control system.
2. Description of the Related Art
A climate control system for a vehicle incorporates a compressor to compress a refrigerant gas. One typical refrigerant compressor for use in a vehicle climate control system is a conventional variable displacement refrigerant compressor, which is provided with a drive shaft driven for a variable rotation about an axis of rotation thereof by an external drive source, pistons slidably fitted in cylinder bores formed in a cylinder block so as to be reciprocated to suck a refrigerant gas from a suction chamber, to compress the sucked refrigerant gas in the cylinder bores, and to discharge the compressed refrigerant gas from the cylinder bores into a discharge chamber, a variable-inclination cam plate mounted to rotate with the drive shaft within a crank chamber and to be operatively engaged with the pistons to cause the reciprocation of the pistons in response to the rotation thereof while reducing the stroke of reciprocating movement of the pistons in response to an increase in a pressure prevailing in the crank chamber, a controlling passage extending between the discharge chamber and the crank chamber to control the pressure in the crank chamber, and a flow control valve arranged in the controlling passage to control the size of an opening in a portion of the controlling passage.
In the above-described variable displacement refrigerant compressor, when a fluorinated hydrocarbons gas is used as the refrigerant gas, and when the refrigerant compressor is incorporated in a refrigerating system operated under a condition such that a discharge pressure and a suction pressure of the refrigerant gas are always kept below a critical pressure of the refrigerant gas (this type of refrigerating system will be hereinafter referred to as a subcritical-cycle-type refrigerating system), it is possible to adjustably change the displacement of the variable displacement refrigerant compressor by the use of the flow controlling valve as schematically shown in FIG. 14.
Referring to FIG. 14, the conventional flow control valve is constructed so as to be arranged in the controlling passage which extends between the discharge chamber and the crank chamber. The flow control valve is provided with a pressure sensing member 80 moving in response to a detection of a change in a suction pressure Ps, and a valve element 81 connected to the pressure sensing member 80 and movable to adjustably open and close a port 83a of the controlling passage 83 in response to the movement of the pressure sensing member 80. The flow control valve receives the suction pressure Ps at the pressure sensing member 80 and moves the valve element 81 in a direction closing the port 83a of the controlling passage in response to an increase in the suction pressure Ps. Further, the pressure sensing member 80 of the flow control valve constantly receives a pressing force F of a spring 82 (this pressing force F of the spring is determined by design) to urge the valve element 81, via the pressure sensing member 80, in a direction opening the port 83a of the controlling passage 83. The valve element 81 is arranged so as to constantly receive a discharge pressure Pd by which the valve element 81 is urged in a direction to open the portion 83 of the controlling passage.
Thus, the above-mentioned flow control valve acts so that the valve element 81 opens the port 83a of the controlling passage 83 when the suction pressure Ps reduces to a pressure below a predetermined set pressure value (it is referred to as a set suction pressure), in order that the refrigerant gas under a discharge pressure Pd flows from the discharge chamber into the crank chamber through the opened port 83a of the controlling passage 83. As a result, when a pressure Pc in the crank chamber is increased, the cam plate is moved toward a position which reduces the angle of inclination thereof, so that the stroke of reciprocating movement of the pistons is reduced. As a result, the displacement of the compressor is reduced.
In accordance with the above-described arrangement of the flow control valve, the valve element 81 of the flow control valve constantly receives the discharge pressure Pd urging the valve element 81 in a direction to close the port 83a of the controlling passage 83. Therefore, when the spring 82 is set so as to exert a predetermined constant force F, the flow control valve indicates such a control characteristics that the set value of the suction pressure Ps acting on the pressure sensing member 80 may be reduced as the discharge pressure Pd acting on the valve element 81 increases. Namely, the relationship between the discharge pressure Pd and the suction pressure Ps which act in the flow control valve indicates a characteristic curve represented by a straight line sloping down from the left to the right in a rectangular coordinates, as shown in FIG. 15. Thus, when the discharge pressure Pd acting on the valve element 81 increases, the set value of the suction pressure Ps acting on the pressure sensing member 80 decreases.
When an actual pressure level of the suction pressure Ps prevailing in the suction pressure region in the refrigerant compressor reduces to a value in an area below the line in FIG. 15, the valve element 81 of the flow control valve is moved to a position opening the port 83a permitting the refrigerant gas under the discharge pressure Pd to enter the crank chamber, and accordingly, when the pressure Pc in the crank chamber is increased, the cam plate is moved to reduce the displacement of the compressor.
Nevertheless, when a refrigerant compressor incorporating therein the above-described flow control valve is operated under a high rotating speed, and when an amount of the refrigerant circulating through a refrigerating system is increased until an excessive increase in the refrigerating performance of the refrigerating system occurs, it is very difficult to quickly reduce the refrigerating performance of the refrigerating system by adjustably controlling the displacement of the refrigerant compressor. This difficulty in controlling the displacement of the compressor is specifically encountered by a refrigerating system of the type in which a closed refrigerant-circulation path of the refrigerating system includes a high-pressure path in which the refrigerant is under a high discharge pressure and, more specifically, is under a supercritical pressure. This type of refrigerating system will be hereinafter referred to as a supercritical-cycle-refrigerating system and, in this system, when the rotating speed of the refrigerant compressor accommodated in the system is increased, the pressure (the discharge pressure) in the high-pressure path can be quickly increased. However, in a low-pressure path of the refrigerant-circulating path, an evaporating pressure (a suction pressure) of the refrigerant cannot be quickly reduced. Thus, when the flow control valve incorporated in the refrigerant compressor has the aforementioned operating characteristics having a straight line relationship between Pd and Ps, and when the rotating speed of the compressor is increased to increase the discharge pressure Pd, the set pressure value of the suction pressure Ps acting on the pressure sensing member 80 of the flow control valve is accordingly reduced to make it difficult to quickly move the valve element 81 in a direction opening the port 83a of the controlling passage 83. Namely, the control of the displacement of the refrigerant compressor is delayed.
EP-0604417B1, based on PCT/N091/00119 (the corresponding published Japanese Translation No. 6-510111), discloses a typical supercritical cycle type refrigerating system including a refrigerant compressor, a heat-radiating type heat exchanger (a gas cooler), a throttling means, a heat-absorbing type heat exchanger (an evaporator), and a liquid-gas separator (an accumulator) which are connected in series to form a closed refrigerant-path. In the disclosed refrigerating system, a temperature at outlet of the gas cooler arranged in the high-pressure path is detected by a temperature sensor, and the operation of the throttling means disposed downstream of the gas cooler in the high-pressure path is controlled on the basis of the detected temperature of the gas cooler outlet to thereby adjust the pressure level prevailing in the high pressure path so that an energy consumption in the refrigerating system is suppressed.
In order to suppress the energy consumption in the supercritical-cycle-type refrigerating system to the minimum, the compressor should be operated under a condition such that a coefficient of performance (COP=Q/W) defined as a ratio of a refrigerating performance (Q) of the evaporator against a compressing work (W) externally applied to the refrigerant compressor becomes the possible maximum value.
It will be understood that the larger a change in the refrigerating performance (Q) of the evaporator is, that is to say, the larger a change in an enthalpy (a difference between an enthalpy at the outlet and that at the inlet of the evaporator) which occurs during the flowing of the refrigerant through the inside of the evaporator is, and the smaller the compressing work (W) necessary for compressing the refrigerant in the refrigerant compressor is, the larger is above-mentioned coefficient of performance (COP) of the refrigerating system. Thus, in the supercritical-cycle-type refrigerating system, when the temperature of the refrigerant detected at the outlet of the heat-radiating type heat exchanger (the gas cooler) in the high-pressure path is kept substantially constant, the coefficient of performance (COP) of the refrigerating system can be increased by increasing a pressure in the high-pressure path to thereby increase the refrigerating performance (Q). This capability of increasing the coefficient of performance (COP) of the supercritical-cycle-type refrigerating system is a remarkable characteristics that could not be exhibited by the subcritical-cycle-type refrigerating system, and accordingly, the operation of the throttling means in the supercritical-cycle-type refrigerating system is different from that of the throttling means included in the subcritical-cycle-type refrigerating system. More specifically, when referring to FIG. 16, which shows a diagram indicating a relationship between a pressure and an enthalpy (a Pressure-enthalpy (P-H) diagram or a Mollier diagram) of a supercritical-cycle-type refrigerating system employing carbon dioxide (CO.sub.2) gas as a refrigerant, it can be seen that the refrigerating performance (Q) of the evaporator is increased in response to an increase in a differential (.DELTA.H.sub.1 =H.sub.A -H.sub.D) between an enthalpy (H.sub.D) at the inlet (the point D) and that (H.sub.A) at the outlet (the point A) of the evaporator, and in response to an increase in an amount of mass flow of the refrigerant flowing through the evaporator. At this stage, when an excessive heating at the outlet (A) of the evaporator increases to an unusually great extent, the specific volume of the refrigerant sucked into the refrigerant compressor increases and the volumetric efficiency of the compressor is reduced in response to an increase in the temperature of the discharged refrigerant, and as a result, an amount of circulation of the refrigerant, i.e., an amount of refrigerant supplied to the evaporator as per a unit time (Kg/h) is reduced to result in an reduction in the refrigerating performance (Q) of the evaporator. Therefore, in order to avoid such a reduction in the refrigerating performance, which is caused by the reduction in the amount of circulation of the refrigerant, by maintaining the extent of the excessive heating at the outlet of the evaporator substantially constant, it is necessary to maintain the enthalpy (H.sub.A) at the outlet (the point A) of the evaporator substantially constant.
On the other hand, the enthalpy (H.sub.D) of the inlet (the point D) of the evaporator is equal to the enthalpy (H.sub.C) at the outlet (the point C) of the gas cooler due to the fact that an expanding process in the throttle means is conducted as an isoenthalpy change. Therefore, the differential (.DELTA.H.sub.1 =H.sub.A -H.sub.D) between the enthalpy (H.sub.D) at the inlet (the point D) and that (H.sub.A) at the outlet (the point A) of the evaporator, and in turn the refrigerating performance (Q) of the evaporator can be increased by reducing the enthalpy (H.sub.C) at the outlet (the point C) of the gas cooler. The interior of the gas cooler in the high-pressure path in which the refrigerant under an supercritical pressure flows, is kept as a single vapor phase occupied by only a high pressure vapor, a pressure in the high-pressure path can be adjusted irrespective of the temperature of the refrigerant at the outlet (point C) of the gas cooler. When the temperature of the refrigerant at the outlet (the point C) of the gas cooler is kept substantially constant, for example, at 40.degree. C., it will be understood from the isothermal line at 40.degree. C. of the P-H diagram of FIG. 16 that higher the pressure in the high-pressure path is, smaller the enthalpy (H.sub.C) at the outlet (the point C) of the gas cooler is. Accordingly, when the temperature of the refrigerant at the outlet (the point C) of the gas cooler is maintained substantially constant, the above-mentioned refrigerating performance (Q (=.DELTA.H.sub.1)), and in turn the coefficient of performance (COP) can be increased by increasing the pressure in the high-pressure path to thereby reduce the enthalpy (H.sub.C) at the outlet (the point C) of the gas cooler. It should be noted that the temperature of the refrigerant at the outlet (the point C) of the gas cooler is substantially equal to the temperature of the air conducting a heat exchange with the refrigerant in the gas cooler.
On the other hand, when the temperature of the refrigerant at the outlet (the point C) of the gas cooler is maintained substantially constant, e.g., at 40.degree. C., and when the pressure in the high-pressure path is increased, the compressing work (W=.DELTA.H.sub.2 =H.sub.B -H.sub.A) to be done by the refrigerant compressor increases.
In this case, since the compression of the refrigerant performed by the compressor is considered to be an adiabatic compression, the compressing operation is processed as an isoenthalpy change, and the compressing work (W) is considered to be equal to a differential between the enthalpy (H.sub.A) at the suction inlet (the point A) of the compressor and the enthalpy (H.sub.B) at the delivery outlet (the point B) of the compressor. Therefore, when the pressure in the high-pressure path is excessively increased, an increase in the compressing work (W) performed by the compressor occurs causing a reduction in the coefficient of performance (COP) of the refrigerating system.
Thus, when the temperature of the refrigerant detected at the outlet (the point C) of the gas cooler is a given temperature, there correspondingly exists a pressure in the high-pressure path which can be determined by the relationship between the refrigerating performance (Q) and the compressing work (W) to be optimum for obtaining the maximum value of the above-mentioned. Therefore, with respect to various temperatures of the refrigerant at the outlet (the point C) of the gas cooler, there are corresponding pressures in the high-pressure path, and accordingly, when the respective pressures are plotted on the P-H diagram, it is possible to obtain an optimum line of control as shown in FIG. 16.
In the supercritical-cycle-type refrigerating system disclosed in EP-0604417B1, the temperature and the pressure of the refrigerant at the outlet (the point C) of the gas cooler are detected by respective sensors, and on the basis of the aforementioned optimum line of control, determination of an optimum pressure in the high-pressure path is carried out with respect to the detected temperature of the refrigerant. Then, the throttle means is controlled so as to adjustably change an actual pressure in the high-pressure path to the determined optimum pressure, and accordingly, the coefficient of performance (COP) of the refrigerating system is increased to the maximum while the energy consumption in the refrigerating system is reduced to the minimum.
In the case of a vehicle refrigerating system, a refrigerant compressor incorporated in the system is driven by a vehicle engine. Therefore, when the speed of rotation of the vehicle engine increases, the drive power applied from the vehicle engine to the compressor is in turn increased. Therefore, an amount of circulation of the refrigerant (Kg/h) flowing through the evaporator is increased, and the refrigerating performance (Q) often becomes excessive. Therefore, in order to prevent the excessive refrigerating performance of the refrigerating system when the number of rotation of the vehicle engine increases, it is necessary to reduce the path of the throttling means, so that the above-mentioned amount of circulation of the refrigerant is reduced. However, when the path of the throttling means is simply reduced, the pressure in the evaporator is reduced to cause a reduction in the temperature of the refrigerant to a saturation temperature corresponding to the reduced pressure, and the required prevention of the excessive refrigeration cannot be achieved. Therefore, when the speed of rotation of the vehicle engine is increased, the size of opening of the throttling means is reduced while simultaneously the displacement of the compressor is correspondingly reduced. Namely, a variable displacement refrigerant compressor which can change its displacement on the basis of detection of a suction pressure (a pressure of the refrigerant at the outlet of the evaporator) and a temperature of the refrigerant at the outlet of the evaporator, is employed so as to reduce the displacement of the compressor, in response to an increase in the number of rotation of the vehicle engine. Thus, the amount of circulation of the refrigerant is reduced in response to the reduction in the displacement of the compressor, and also the temperature of the refrigerant in the evaporator due to an increase in the suction pressure, i.e., an increase in the pressure of the refrigerant in the evaporator caused by the reduction in the displacement of the compressor can be obtained. Consequently, excessive refrigeration due to an increase in the speed of rotation of the vehicle engine can be effectively prevented.
Nevertheless, in the above-described supercritical-cycle-type refrigerating system, when the flow control valve described with reference to FIGS. 14 and 15 is incorporated in a variable displacement compressor to adjustably change the displacement thereof, there occurs a problem such that the control of the displacement of the compressor in response to a change in an increase in the speed of rotation of the vehicle engine cannot be quickly achieved during the supercritical refrigerating cycle. Namely, in the supercritical refrigerating system, the temperature and the pressure of the refrigerant at the outlet (the point C) of the gas cooler in the high-pressure path are detected, and the throttling means is regulated so that the pressure of the refrigerant at the outlet (the point C) of the gas cooler is varied to an optimum pressure corresponding to the detected temperature, and as a result, the maximum coefficient of performance (COP) and in turn, the minimum energy consumption of the supercritical refrigerating system are achieved. In the supercritical refrigerating system for a vehicle, which requires a regulating operation of the throttling means, when an increase in the speed of rotation of the vehicle engine and in turn the rotating speed of the drive shaft of the refrigerant compressor (the variable displacement compressor) occur, a mass amount of the refrigerant supplied to the gas cooler is increased. Thus, a pressure of the refrigerant in the gas cooler (i.e., a pressure in the high-pressure path and a discharge pressure of the compressor) is increased. Further, as described hereinabove, since the throttling means is regulated so that the pressure at the outlet of the gas cooler is kept substantially constant, the path of the throttling means must be increased to prevent an increase in the pressure at the outlet of the gas cooler. Therefore, the operation of the throttling means to reduce the path thereof is often slow to result in that the control of the refrigerating performance cannot be quickly achieved.
As will be understood, in accordance with an operating characteristics of the supercritical-cycle-type refrigerating system, when the number of rotation of the drive shaft of a compressor incorporated in the system is increased, a pressure in the high-pressure path of the refrigerating system, i.e., a discharge pressure of the refrigerant delivered by the compressor can be quickly increased, but a pressure in the low-pressure path, i.e., a suction pressure to be sucked into the compressor cannot be quickly reduced. Therefore, when the flow control valve as described in connection with FIGS. 14 and 15 is incorporated in the compressor, the set value of the suction pressure (Ps) acting on the pressure sensing member of the valve is reduced by an increase in the pressure prevailing in the high-pressure path, and accordingly, an occurrence of an excessive refrigeration of the system cannot be successfully prevented.